Multistage compressor

ABSTRACT

The invention relates to a multistage compressor for providing high-pressure gas in filling stations, said compressor consisting of a high-pressure compressor and a multistage booster compressor (W). Both compressors comprise membrane pump chambers which are controlled by camshafts. The booster compressor (VV) can contain a plurality of stages (S 1 , S 2 , S 3 ), the chambers in each stage forming groups (G 1 , G 2 , G 3 ) of chambers. The chambers of a group are synchronously operated in phase. The chambers of two successive stages are operated in opposition of phase. The number of chambers in a group is approximately the same as the pressure ratio π, so that the size of the chambers can be standardised.

The invention refers to a multistage compressor for compressing gases and in particular to a membrane compressor for a gas filling system for fueling a motor vehicle operating on natural gas, methane or similar gases as well as on hydrogen.

Gases used as a means of energy storage in a motor vehicle are problematic because of the required storage volume that, for natural gas, is larger by 3 orders of magnitude under ambient conditions, compared to liquid energy carriers. For this reason, it has been regulated that natural gas is available at filling stations at a pressure of 250 bar, so that a pressure of 200 bar, as defined by technical rules, is reached and not exceeded in a pressure gas container of a vehicle at a reference temperature of 15° C. Thus, only about three times the storage volume of a petrol vehicle have to be made available in the car.

In gas filling stations for direct filling with a compressor, the compression causes an undesirable heating of the gas, whose effect is the greater, the greater the stage pressure ratio X is. To achieve the desired end pressure, the pressure ratio π may be reduced by increasing the number of stages.

In gas filling stations, the pressing work to be done causes a heating of the gas in the pressure gas container. The Joule-Thomson effect (change in the gas temperature by throttling) of the real gas generally counteracts this heating. However, it is only under very favorable conditions, i.e. at sufficiently low temperatures, that the Joule-Thomson effect and the heat dissipation to the environment will be sufficient to compensate for the heating caused by the work of pressing the gas. If these favorable conditions do not exist, a short-filling of the pressure gas container will occur in gas filling stations without a cooling device during decanting or direct filling. This is due to the fact that the pressing work causes a high temperature and thus a correspondingly high-pressure in the pressure gas container, whereby the differential pressure available for filling is reduced so much that the filling operation takes long and is therefore interrupted before the pressure gas container holds the gas volume possible by technical specifications.

DE 197 05 601 A1 describes a natural gas filling method without a cooling of the gas, wherein the operation of filling the pressure gas container is carried on until the pressure in the line to the pressure gas container exceeds a maximum pressure. Another possibility provides that the filling operation is interrupted when the mass flow exceeds a threshold value.

WO 97/06383 A1 describes a gas charging system for high-pressure gas bottles. In this case, the gas is cooled by flushing the high-pressure gas bottle to be filled, whereby two connectors for the feed and return lines are required. In the flushing circuit, the gas is cooled by a heat exchanger or by being mixed with the gas in the supply container.

EP 0 653 585 A1 proposes a system for filling a pressure gas container. It describes the execution of a testing impulse which is evaluated using the thermal equation of state for the real gas. Further, a switching to supply containers with higher pressure (multi-bank method) during the filling is described. The filling operation is executed intermittently. No cooling device for the gas is provided.

DE 102 18 678 B4 describes a method and a device wherein the gas to be filled into the pressure gas container is fed from a high-pressure supply container via a cyclone tube as a cooling device. The cyclone tube utilizes the existing differential pressure in the filling system to separate the gas flow into a hot gas flow and a cold gas flow. The latter will then be fed to the pressure gas container.

DE 10 2005 006 751 A1 also proposes a device, wherein a decrease in the temperature of gases is obtained through a cyclone tube without providing for a separation into a hot gas flow and a cold gas flow.

The operation of the two latter methods and devices is based on the fact that the gas is fed to a swirling device at a supercritical pressure ratio, which swirling device is arranged axially between two pipes that have different inlet diameters. A decrease in the temperature of gases by means of a cyclone tube will be successful if, and only if, supercritical pressure conditions prevail. At a critical pressure ratio for natural gas of 1/π′<0.5427 and a pressure in the supply container of p_(D)=250 bar, which is typically under-run when a plurality of vehicles are filled in short succession, a subcritical state is reached when the pressure in the pressure gas container has risen to P_(D)=135 bar. This means that under the conditions set by the technical specifications, the use of a cyclone tube will provide no further decrease in the temperature of the gas when a pressure gas container is filled with natural gas in the pressure range from P_(D)=135 bar to p_(D)=200 bar.

In WO 01/27475 A1, a multistage membrane compressor of a star-shaped structure is described, wherein the compressor chambers are arranged in a star shape around a crankshaft. The compressor chambers form individual stages of a multistage compressor and therefore have different volumes. Thus, high compression ratios can be realized, however, while producing substantial amounts of compression heat.

Direct filling independent of a filling station is feasible where installing public natural gas filling stations is uneconomical. Vehicles—and not only those of individual transport—could be filled where they are during their downtime. This may be in industrial parks, garages or car boards. Numerous households or buildings have natural gas at their disposal for hating purposes. Using a compressor (natural gas compressor), this natural gas could be compressed in a garage during the night from the typical natural gas network pressure level of 50 mbar to 200 bar at a reference temperature of 15° C. The motor vehicle can be filled therewith.

Another possibility to use such a filling system is envisaged in agriculture, where biological gas is produced in vast amounts. Instead of feeding this biological gas into a public gas network, the gas may be compressed in situ and be used to drive agricultural vehicles and machines. Thus, it would be possible in the future to replace biological diesel fuel in the field of agriculture. One demand, among others, to be met by this compressor is that the compressor has to be designed such that a full filling can be achieved in one night (ca. 8 hours) at 200 bar and a reference temperature of 15° C.

The major problem of a multistage high-pressure compressor is the heating of the gas between the individual compressor stages and the cooling of the gas at the compressor outlet, which must never exceed a temperature of 60° C. when entering the pressure gas container during the filling.

It is an object of the present invention to design a multistage compressor such that it becomes possible to fill a pressure gas container by direct filling such that a limit value of the pressure, given by technical specifications, is reached in the pressure gas container at a predetermined limit temperature.

The multistage compressor of the present invention is defined in claim 1. According to the same, a booster compressor is located upstream of the high-pressure compressor, which booster compressor includes at least one booster stage, wherein each booster stage comprises several chambers that are combined into at least one group and the chambers of a group are operated in common and synchronously.

The invention is based on the idea that, in a high-pressure compressor of membrane structure, a reduction of the pressure ratio should not be obtained by increasing the number of stages of the high-pressure compressor, but that a single- or multistage booster compressor should be provided upstream of the high-pressure compressor.

With such a booster compressor, not only the pressure ratio π can be decreased, but—without changing the dimensions of the high-pressure compressor—also the mass flow rate can be increased proportionally to the increase of the boost pressure. The combination of a booster compressor and a high-pressure compressor—both realized as membrane structures—allows to establish a gas filing system that meets a wide range of demands with respect to the mass flow rate (filling time). This is achieved with a single- or multistage compressor, if the membrane dimensions thereof identically match the dimensions of the membrane of the first stage in the high-pressure compressor. This is true irrespective of the number of stages, if an integral value π=2, 3, 4, . . . is chosen for the pressure ratio in the booster compressor.

The invention allows to use standardized multiple membrane pumps. In this case, the sizes of the different pump chambers may be equal to each other. It is also possible, however, in a pumping device with pump chambers in a star-shaped arrangement, to make the pump chambers replaceable so that pump chambers of different sizes are available which may optionally be mounted in the pumping device.

In direct filling, it becomes possible with a multistage booster compressor to maintain a supercritical pressure ratio until the end of filling of a pressure gas container at a container pressure of 200 bar. As already comprehensively described in DE 10 2005 016 114 A1, this allows to decrease the gas temperature through the use of a cyclone tube or a spray-nozzle element according to DE 100 31 155 C2 or only through adiabatic throttling, so that an increase in temperature caused by the pressing work can be compensated. In this manner, filling at a container pressure of 200 bar at a reference pressure of 15° C. can be realized even in direct filling, regardless of the ambient temperature.

If, according to the invention, the high-pressure compressor is also realized as a membrane compressor, it may be operated at the same speed via a common shaft with the upstream membrane booster compressor.

A particularly suitable embodiment of the invention provides that, with very large gas volume flows, the first and the following stages of the booster compressor are divided into several membrane chambers if the membrane diameter required for a large volume flow can not be realized for material related reasons. For example, in a single-stage booster compressor with a pressure ratio π=2, the first stage of the high-pressure compressor can be supplied with twice the gas volume flow of a high-pressure compressor operated without a compressor. To achieve this, the booster compressor has to be provided with two membrane chambers of the same dimensions as the first stage of the high-pressure compressor. Depending on the boost pressure to be generated in the booster compressor and the gas volume flow increasing proportionally thereto, different variations can be made to the booster compressor design according to the invention, which variations will be explained hereunder using a simple mathematic consideration.

Applying the geometric series

I+x+x²+x^(n)  (1)

as the base, where the quotient of two successive members is constant and, in the present case, the value x describes the pressure ratio π and the exponent n describes the number of membrane chambers per stage, so that the above equation can be represented in the following form:

I+π+π²+π³+π^(n)  (2)

From equation (2), the following is obtained for a pressure ratio of

π=2:1+2+4+8+2^(n)

π=3:1+3+9+27+3^(n)

π=4:1+4+16+64+4^(n)

wherein, in the practical implementation, first, only the combinations listed in the table below may be realized.

pressure number of number of ratio π stages z chambers n 2 1 2 2 4 3 6 3 1 3 2 9 4 1 4

For a pressure ratio of π=2, the dimensions of the two membrane chambers of the last stage of the booster compressor have to be the same as those of the first stage of the downstream high-pressure compressor of membrane structure. For a pressure ratio of π=3, this is true for the three membrane chambers, and for a pressure ratio of π=4, this applies to the four membrane chambers of the last stage of the booster compressor.

Provided that the booster compressor is operated at the same speed of rotation as the downstream high-pressure compressor of the membrane type and the stroke in the individual membrane chambers is the same as the membrane stroke in the first stage of the high-pressure compressor, it can be proven that the diameter of the membrane chamber always corresponds to the membrane diameter of the first stage of the high-pressure compressor. The pressure ratio π must always correspond to the number z of the membrane chambers/stages and be an integral π=2, 3, 4 that is independent of the stage pressure ratio π_(HD) in the high-pressure compressor.

For a single-stage booster compressor with a pressure ratio π=4, a diameter D₁ and

-   -   4 membrane chambers in the 1st stage: z₁=4         it can be shown that     -   1st stage: V₁=z₁·D₁ ²=4D₁ ²     -   HD₁: V_(HD1)=V₁/π=(4D₁ ²)/4         -   D_(HD1)=(D₁ ²)^(0.5)=D₁

Here, V₁ is the volume of the 1st stage, D₁ is the diameter of the 1st stage of the booster compressor, HD₁ is the diameter of the first stage of the high-pressure compressor, V_(HD1) is the volume of the 1^(st) stage of the high-pressure compressor, D_(HD1) is the diameter of the 1^(st) stage of the high-pressure compressor, z_(n) is the number of the membrane pumps in the stage n.

For a two-stage booster with a pressure ratio π=3, a diameter D₁ and

-   -   9 membrane chambers in the 1st stage: z₁=9     -   3 membrane chambers in the 2nd stage: Z₂=3         the following is obtained     -   1st stage: V₁=z₁·D₁ ²=9D₁ ²     -   2^(nd) stage: V₂=V₁/π=(9D₁ ²)/3=3D₁ ²         -   D₂=(3D₁ ²/z₂)^(0.5)=(3D₁ ²/3)^(0.5)=D₁     -   HD₁: V_(HD1)=V₂/π=(3D₁ ²)/3         -   D_(HD1)=(D₁ ²)^(0.5)=D₁             and for a three-stage booster compressor with the pressure             ratio π=2, a diameter D₁, as well as     -   8 membrane chambers in the 1st stage: z₁=8     -   4 membrane chambers in the 2nd stage: z₂=4     -   2 membrane chambers in the 3rd stage: z₃=2         the following is obtained     -   1st stage: V₁=z₁·D₁ ²=8D₁ ²     -   2^(nd) stage: V₂=V₁/π=(8D₁ ²)/2=4D₁ ²         -   D₂=(4D₁ ²/z₂)^(0.5)=(4D₁ ²/4)^(0.5)=D₁     -   3^(rd) stage: V₃=V₂/π=(4D₁ ²)/2=2D₁ ²         -   D₃=(2D₁ ²/z₃)^(0.5)=(2D₁ ²/2)^(0.5)=D₁     -   HD₁: V_(HD1)=V₃/π=(3D₁ ²)/2=D₁ ²         -   D_(HD1)=(D₁ ²)^(0.5)=D₁

It is considered a particular advantage of a booster compressor designed in this manner that like components can be used in a cost-saving manner when subdividing the stages into individual membrane chambers. These are, as essential parts of the booster compressor, the membranes of the membrane chambers which all have the same dimensions as the membrane of the first downstream high-pressure compressor of the membrane type. It is another advantage of the present invention that the dimensions of the membrane or the membrane chambers are independent of the pressure ratio, the volumetric flow rate and the desired output pressure of the booster compressor. According to the invention, a booster compressor further allows to first increase the delivery volume of a high-pressure compressor of the membrane type by two to nine times with respect to an operation without a booster compressor.

The following is a detailed description of embodiments of the invention with reference to the accompanying drawings.

In the Figures:

FIG. 1 is a schematic longitudinal section through a multi-chamber membrane pump with membrane chambers arranged in a star shape around a camshaft,

FIG. 2 illustrates an embodiment of a compressor formed by a booster compressor and a single- or multistage high-pressure compressor of the membrane type,

FIG. 3 is a schematic general view on the structure and the division of the membrane chambers in a compressor with a single-stage booster compressor and a single-stage high-pressure compressor,

FIG. 4 is an embodiment with a two-stage booster compressor and a single-stage high-pressure compressor,

FIG. 5 is an embodiment with a three-stage booster compressor and a single stage high-pressure compressor,

FIG. 6 is a schematic illustration of the camshaft in a single-stage booster compressor,

FIG. 7 is an illustration of the camshaft in a two-stage booster compressor, and

FIG. 8 is an illustration of the camshaft in a three-stage compressor.

FIG. 1 is a schematic illustration of a membrane chamber pump 50. A camshaft 52 is supported in a housing 51, the camshaft controlling a plurality of membrane pump chambers 53 arranged in star shape around the camshaft. Each membrane pump chamber 53 is delimited by a flexible membrane 54 that is movable between two end positions. The membrane pump chamber 53 has an inlet line 54 with a check valve 55 and an outlet line 66 with a check valve 57. gas is drawn in through the inlet line 64 and the compressed gas is expelled through the outlet line 66.

The membrane 54 is moved by a liquid buffer 58 contained in a cylinder 59 in which a piston can be moved. A spring 61 presses the piston 60 against a ball bearing 62 sitting on the camshaft 52 and forming an eccentric for driving the piston. The piston 60 performs a linear movement in the cylinder 59 whereby the membrane 54 is moved between its end positions through the liquid buffer 58.

Besides the ball bearing 62, FIG. 1 illustrates further ball bearings on the camshaft 52. These serve to actuate the other membrane chambers arranged in star shape around the camshaft.

The liquid buffer 58 is supplied or maintained filled with liquid by a pumping device (not illustrated) that is also driven by the camshaft.

FIG. 2 illustrates a compressor formed by a booster compressor VV and a high-pressure compressor HD. The booster compressor is formed by the multiple membrane compressor. It comprises four membrane pump chambers 11, 12, 13, 14 controlled by the same camshaft. The outlet lines of the membrane pump chambers are illustrated in FIG. 2. They are connected with a manifold 70 leading to a membrane pump chamber FD1 of the high-pressure compressor HD. In the present case, the high-pressure compressor is also designed as a camshaft controlled membrane pump, the other membrane pump chambers being adapted for use as further stages of the high-pressure compressor or for other purposes.

In the booster compressor VV, all membrane pump chambers 11-14 are controlled synchronously and in phase. This means that all chambers take in simultaneously and expel the compressed gas in phase. Likewise, the booster compressor VV and the high-pressure HD are synchronized with each other, the chamber HD1 that receives the compressed gas from the booster compressor VV being operated in opposite phase relative to the booster compressor. In other words: when the booster compressor expels gas, the chamber HD1 has to be in a position for receiving gas.

In the present embodiment, it is assumed that the booster compressor has a compression ratio of π=4. This means that the entire gas received by the booster compressor is compressed in the booster compressor to one fourth of its volume. Since, on the other hand, the volumes of four chambers are combined, a gas volume is obtained that, in the compressed state, approximately corresponds to the volume of one of the chambers of the booster compressor. Thus, the chamber HD1 of the high-pressure compressor has approximately the same volume as one of the chambers of the booster compressor.

FIG. 3 illustrates the structure of a single-stage booster compressor with four membrane chambers, each having a pressure ratio of π=4 and an output pressure p_(A)=4 bar. The first figure in the schematically illustrated membrane chambers refers to the compressor stage, the second figure sequentially numbers the membrane chambers in the associated stage. HD1 refers to the first stage of the downstream high-pressure compressor, whose dimensions are identical with the individual membrane chambers of the booster compressor.

FIG. 4 illustrates a corresponding structure for a two-stage booster pressure with a pressure ratio π=3 and an output pressure p_(A)=9 bar, wherein a total of twelve membrane chambers are implemented in the booster compressor. Finally, FIG. 5 illustrates the structure of a three-stage booster compressor of the membrane type. Here, a pressure ratio of π=2 was chosen, whereby an output pressure p_(A)=8 bar is obtained in the third stage with a total of fourteen membrane chambers.

In the embodiment of FIG. 3, the booster compressor VV is formed by a single stage S₁ with n=4 chambers, the pressure ratio being π=4. In the embodiment of FIG. 4, the booster compressor VV is formed by two stages S₁, S₂, each stage having n=3 chambers combined into a group G₁-G₃. In the embodiment of FIG. 5, the booster compressor VV is formed by three stages S₁, S₂, S₃, where four groups of chambers are formed in the first stage S₁ and each group is formed by n=2 chambers. The compression ratio π is also 2. In the second stage as well, two respective chambers are combined into a group, and the same is true for the third stage S₃.

FIG. 6 is a cross section through and a view of the camshaft for a single-stage booster pressure. By the eccentric shape of the camshaft that extends over a range from 0° to 180°, as is evident from the cross section, the individual membranes in the membrane chambers of the booster compressor are operated during one half of a rotation of the camshaft. The membrane chambers operated in a two-stroke mode take in the gas in the first quarter up to the upper dead centre at 90° and compress the same in order to subsequently begin the expulsion process in the second quarter, which is completed at an cam angle of 180°. As is evident from the illustration, in a single-stage booster compressor with four membrane chambers, the entire gas volume flows into the first stage of the downstream high-pressure compressor. In the illustration, the first figure in the schematically shown eccentric of the camshaft refers to the compressor stage in the booster compressor, whereas the second figure indicates the sequential number of the membrane chambers present in the respective compressor stage.

FIG. 7 is a cross section through and a view of the camshaft using the example of a two-stage booster compressor. However, in the first stage, the intake and the compression of the gas starts at an angular position of the cam of 180°, the upper dead centre being at a cam angle of 270° and the expulsion into the second stage of the booster compressor is completed a cam angle of 0°. The timing of the intake, the compression and the expulsion of the gas into the first stage of the downstream high-pressure compressor in the second stage will then be the same as in a single-stage booster compressor.

FIG. 8 is a cross section and a view, illustrating the course of the control of the membranes in the membrane chambers by a camshaft for a three-stage booster compressor with a low stage pressure ratio of π=2. The compression process in the first stage starts at an angular position of the cam of 0°, is completed at the upper dead centre at a cam angle of 90°, followed by the process of expulsion into the second compressor stage that is completed at a cam angle of 180°. The second compressor stage has its upper dead centre at a cam angle of 270°, followed by the process of expulsion into the third compressor stage up to a cam angle of 0°. The latter has its upper dead centre at a cam angle of 90° and ends with the subsequent expulsion of the gas flow from the two membrane chambers of the first stage of the booster compressor into the first stage of a downstream high-pressure compressor. This process is completed at a cam angle of 180°.

The gas forces to be controlled by the camshaft in the form of an eccentric have a balanced mass moment of inertia only in the two-stage embodiment of the booster compressor (FIG. 7). This means the mass moment of inertia caused by the gas forces has to be balanced only for even numbers of compressor stages. For odd numbers of stages (FIGS. 6 and 8), the gas forces of a stage that cause the mass moment of inertia that does not have to be balanced must be balanced by a circumferential counter-weight at the camshaft. 

1. A multistage compressor comprising a high-pressure compressor (HD) having at least one periodically driven membrane pump chamber (HD1), characterized in that a booster compressor (VV) is provided upstream of the high-pressure compressor (HD), said booster compressor having at least one booster compressor stage (S₁-S₃), each booster compressor stage comprising a plurality of chambers (11-14) combined into at least one group (G₁-G₄), and that the chambers of a group are driven together and synchronously.
 2. The multistage compressor of claim 1, characterized in that the number (n) of the chambers in a group (G₁-G₄) of the booster compressor (VV) is substantially equal to the compression ratio (π) of this stage.
 3. The multistage compressor of claim 1, characterized in that the booster compressor (VV) includes at least two booster compressor stages (S₁-S₃), each chamber of the second booster compressor stage (S2) being fed by a group (G₁-G₄) of chambers of the first booster compressor stage (S₁).
 4. The multistage compressor of claim 3, characterized in that a third booster compressor stage (S₃), where each chamber is fed by a group of chambers of the second booster compressor stage (S₂).
 5. The multistage compressor of claim 1, characterized in that the chambers are operated by one or a plurality of synchronized camshafts (52), and that, according two the two-stroke method, an intake and a compression of the gas is performed in the chambers (53) during one half of a rotation of a camshaft between 0° and 180°, which is completed at an upper dead centre at a cam angle of 90° and which is followed by an expulsion of the gas into the next stage up to a camshaft angle of 180°.
 6. The multiple compressor of claim 1, characterized in that, with several chambers existing in a stage of the booster compressor (VV), the compression of the gas is performed synchronized in time and, when the compression is completed, the gas is delivered to two or more chambers of the following stage of the booster compressor also synchronized in time.
 7. The multistage compressor of claim 1 characterized in that the gas leaving two or more chambers of the last stage of the booster compressor is introduced into the first stage of a high-pressure compressor (HD) at the end of the compression stroke.
 8. A multistage compressor of claim 1, characterized in that the dimensions of the membrane chambers (53) of the booster compressor (VV) are configured such that membranes (52) of equal diameter may be used for different structural sizes of a membrane compressor that cover a range of the multiple pressure increase and gas delivery.
 9. The multistage compressor of claim 1, characterized in that the mass moment of inertia caused by the gas forces is balanced in an even number of stages of the booster device (VV).
 10. The multistage compressor of claim 1, characterized in that, with an odd number of stages of the booster compressor (VV), the mass moment of inertia, which is caused by the gas forces in a stage and is not balanced, is balanced by a circumferential counter weight at the camshaft. 